Impact wrench



Oct. 29, 1940. M. HUTCHISONQ JR 2,219,869

' IMPACT WRENGI-l Filed March 19, 1.925? 6 Shets-Shet l l\\\ Wan 4:

BY I

ATTORNEY Oct. 29, 19400 M. R. HUTCHISON. JR 2,219,369

IMPACT WRENCH 6 Sheets-Sheet 2. I

Filed March 19, 1937 ATTORNEY Oct. 29, 1940. R. HUTCHISON. 4R 2,219,869

IMPACT WRENCH Filed March 19, 1937 6 Sheets-Sheet 5 ATTORNEY 5' 3 IMPACT WRENCH Filed March 19, 1957 6 Sheets-Sheet 4 Hm \Nwm F ii:

NM Nw mm Oct; 29,1940. M. HUTCHISON. JR 2,219,869.,

IMPACT WRENCH Filed March 19. 1937 6 Sheets-Sheet 6 ATTORNEY Patented Oct. 29, 1940 UNITED STATES IMPACT WRENCH Miller R. Hutchison, Jr., Madison, N. J., assignor to Chicago Pneumatic Tool Company, New York, N. Y., a corporation of New Jersey Application March 19, 1937, Serial No. 131,830

12 Claims.

This invention relates to power transmission systems for converting continuous torque into pulsating torque and has particular reference to power operated wrenches adapted to drive nuts, bolts and the like with a rotary hammer action.

Conventional types of impact wrenches comprise: a continuously rotatable motor; an intermittently rotatable hammer or impacting element; and some kind of flexible coupling permitting the motor to drive the hammer without lost motion. As a consequence of the use of a flexible coupling, the hammer intermittently acquires a rotative speed exceeding that at which it would be driven if infiexibly connected to the driving motor. This phenomenon attends every use of a torsionally flexible coupling in machinery between two rotating parts not always run ning at a common speed.

Earlier wrenches of the accumulator or impact type employed separable clutch teeth between the hammer and the anvil driven thereby. This clutch arrangement served to transmit the applied torque (at the driving and of the flexible coupling) until the anvil encountered a predetermined resistance to rotation. Then, the hammer and anvil were forcibly declutched to permit the hammer to accelerate under the influence ofthe flexible coupling and subsequently reengage with an impact.

The remeshing of the clutch elements in such prior designs was random and success depended upon correctly anticipating hammer position as a function of angular displacement of the flexible coupling. Accordingly, much emphasis was placed upon making disengagement of the clutch to be dependent upon a predetermined resistance to rotation of the anvil, since this was a function of the energy stored in the flexible coupling or accumulator and the subsequent transfer of energy and the acceleration of the hammer during its free movement.

As a result of the need to anticipate the relative positions of the hammer and anvil during the acceleration period, these wrenches could be operated successfully only within a relatively narrow range of speeds, limiting their performance to a restricted class of work. Operation at speeds outside this rangeresulted in the hammer (a) striking the anvil endwise before the impact faces of the clutch had been oriented into their proper phase relation, or (b) striking the anvil rotationally with only partial meshing of the clutch teeth. Improper reengagement of hammer and anvil in these ways either: (a) reduced the angular velocity of the hammer before the rotary blow was struck, thereby attenuating the rotary impact; or (b) caused rapid deterioration of-the clutch teeth due to excessive stresses over small portions of the impact areas provided. I

The random timing characteristics of these earlier impact wrenches seriously limited the designer of the flexible coupling to a few forms and arrangements.

Among the objects of the present invention are:

1. To guide the movement of the hammer element of the impact clutch so that the clutch teeth are completely meshed at the instant of impact, irrespective of the speed of rotation, thereby adapting the wrench for operation at all rotative speeds.

2. To make it possible to apportion the force of the rotative impact to the work required to be done by governing the torque or speed of the driving motor. The intensity of the blow becomes a function of the motor torque or square of the number of motor revolutions per minute, respectively.

3.To permit selection of a driving connection between the motor and hammer in which the torsional'rate is unimportant.

This invention makes it possible to employ any sort of coupling between the motor and hammer which is of suflicient strength to sustain the torque applied to it, such coupling being either rigid or extremely soft, or any value between these limits, without materially affecting the operation of the tool or requiring its operation at any particular speed in relation to the torsional rate of the coupling.

According to one feature of the present invention, the hammer is grooved to provide a track for'guiding its movement around the anvil. Conversely, the anvil may be grooved to provide a track for the hammer. Thus, relative angular displacement is always accompanied by a corresponding relative axial movement, irrespective of the speed or direction of rotationor applied torque. The hammer is driven by a lead screw imparting an axial component of force tending to declutch the impact shoulders between the hammer and anvil. Following disengagement of the shoulders the hammeris rorelative positions of the hammer and anvil at the instant of impact.

Another object of the invention is the elimination of lost motion between the rotatable impact hammer and the anvil which it drives. In accordance with this object, the hammer and anvil are so constructed and arranged that torque is delivered to the anvil during the interval between impacts. An advantage of this arrangement is that when the hammer blow is delivered to the anvil there is no slack in the driving connection between the anvil and the driven nut and bolt, which otherwise would attenuate the force of the blow.

A further object of the invention is the provision of a resilient coupling which provides lost motion when a predetermined torque is exceeded.

Still further objects are the provision of a. resilient coupling in a power transmission system, which is light in weight, compact, durable, and easy to service and whose resiliency may be varied at will. According to a feature of this invention, the resilient coupling forms a part of an epicyclic speed reduction unit and comprises an air-operated piston and cylinder interposed between the floating member and the stationary one. Certain of the gears rotate about a crank shaft connected to the air-operated piston and adapted to oscillate to provide resiliency. According to a further feature of the invention said crank shaft is arranged to revolve through a complete revolution to provide lost motion if the resistance to rotation becomes excessive.

Still another object of the invention is the provision of a governor for a rotary air motor which is simple in construction and which may be readily adjusted for any desired speed.

Other objects and features of the invention will appear more clearly from the following description taken in connection with the accompanying drawings and appended claims.

In the accompanying drawings wherein is illustrated one embodiment of the invention:

Fig. 1 is a longitudinal section of the front portion of a power-operated wrench, illustrating the hammer, anvil and lead screw mechanism;

Fig. 2' is a front end elevation of the structure shown in Fig. 1;

Fig. 3 is a cross section as indicated by the arrows 3 in Fig. 1, the outer casing being omitted;

Figs. 4 and 5 are crosssections as indicated by the arrows 4 and 5 respectively in Fig. 1;

Fig. 6 is alongitudinal section of a rotatable hammer assembly, as indicated by the arrows 6 in Fig. '7, and showing one of two identical halves of the hammer;

Fig. '7 is a front end elevation of the hammer assembly;

Fig. 8 is a cross section of the anvil, looking rearwardly as indicated by the arrows 8 in Fig. 9;

Fig. 9 is a side elevation of a fragmentary portion of the anvil;

Figs. 10 and 11 are cross sections of the hammer, as indicated by the arrows i0 and H respectively in Fig. 6;

Fig. 12 is a development of the inside of the hammer, showing the positionof the anvil with relation to the internal groove in the hammer;

Figs. 13, 14 and 15' are fragmentary developments similar to Fig. 12 but showing the hammer in difierent positions relative to the anvil;

Fig. 16 is a perspective view of the lead screw, hammer and anvil assembly, an annular sector of the hammer being cut away to show the internal groove;

Fig. 17 is a perspective view of the lead screw;

Fig. 18 is a continuation of Fig. 1 showing, in longitudinal section, the prime mover and power transmission for driving the mechanism shown in Fig. 1;

Fig. 19 is an elevation of a part of the governor shown in Fig. 18;-

Fig. 20 is a front elevation of the governor element shown in Fig. 19;

Fig. 21 is a cross section through the piston and cylinder of the resilient coupling, as indicated by the arrows 2| in Fig. 18, the piston being shown in the no load position;

Fig. 22 is a view similar to Fig. 21 showing, in full lines, the resilient coupling under stress and, in broken lines, one of the positions of the piston during the lost motion period;

Fig. 23 is a graph showing the relation between the torque transmitted by the resilient coupling and the deflection of the crank shaft away from its no-load position;

Fig. 24. is a cross section through the rotary motor and reversing valve, looking rearwardly, as indicated by the arrows 24 in Fig. 18;

Fig. 25 is a cross section through the governor, looking forwardly, as indicated by the arrows 25 in Fig. 18, and

Fig. 26 is an enlarged longitudinal section of the governor showing the segments of the collar thrown outwardly by centrifugal force.

Referring particularly to Fig. 1, the hammer and anvil assembly is driven by a rotatable drive shaft 20. Any suitable means may be employed for driving the shaft, a preferred form of which will be described later. Shaft 20 is supported for rotation without axial movement by means which include a roller thrust bearing 2|. The thrust bearing is supported by a bushing 22 and is held between an inwardly extending flange on the bushing and a retaining plate 23. The bushing 22 is mounted in a bearing support housing 24, as shown in Figs. 1 and 5. Bolts 25, passing through openings in the retaining plate and in the housing 24, are threaded to ears on the bushing. A front casing 26 seats against an annular shoulder on the bearing support housing 24 and is secured by bolts 21 shown in Fig. 2. The rear A lead screw. sleeve 30 surrounds the drive shaft 20 and is provided with internal splines fitting a fluted portion 3| on the shaft. Sleeve 30 seats at its rear end against thrust bearing 2| and is held by a retaining nut 32 screwed on a reduced extension 33 of the shaft 20. A look tab Washer 34 engaging flutes 3| prevents nut 32 from turning relative to shaft 20. Thus, the drive shaft 20, lead screw sleeve 30 and retainer nut 32 rotate as a unit and are held against axial movement. 7

The front end of the casing 26 supports a cast insert 35 in which a thrust bushing 36 is held with a press fit. A tool head 38, having an interior recess 38a, is rotatably mounted in bushing 36. The rear end of the tool head is rotatably seated against the retainer nut 32. A flange 38b on the tool head seats against bushing 36. Bushing 36 and nut 32 thus cooperate to prevent ax ial movement of the tool head. The extension 33 on the shaft 20 carries a roller bearing 39 which fits the recess 38a and which cooperates with the nut 32 to hold the tool head 38 and drive shaft 20 in axial alignment. The front extremity of the tool head consists of a square portion 40 projecting from casing 26 and adapted to be connected to a wrench socket (not shown). An oil seal 4| may be provided on the tool head 38 in front of bushing 36.

The means for transmitting power from the drive shaft 20 to the tool head 38 comprises a rotatable impact hammer 43 and an anvil 44, the latter being integral with the tool head. As shown in Figs. 1, 8, 9 and 16, the anvil comprises a pair of segmental wings a and 'b radiating from a hub portion. The wings, which are identical in shape, are 90 in circumferential extent and spaced 90 apart. Each wing has a cylindrical periphery, approximately radial edges, and helical side faces. The hammer comprises two identical halves 43a and 43b held together by transverse bolts 45 and aligned by dowel pins 46. The front end of the hammer is supported for relative rotary and axial movements on anvil 44, by means of a split bushing 41. Dowel pins 48, pressed into radial holes in the hammer, secure the bushing to the latter.

' A nut is provided with internal threads complementary to the threads on the lead screw sleeve 30. The nut fits a portionof an, axial bore in the hammer 43 and has a flange engaging the rear end of the hammer. Bolts 52, passing through the flange, secure the nut 50 rigidly to the hammer. A wire 53 (see Fig. 4) passes through the heads of the bolts and serves as locking means therefor.

The interior of the hammer is formed with two spiral grooves 55, each 270 in length and overlapping the other groove over an arc of 90. The width'of each groove is equal to, or slightly greater than, the width of the wings a and b on the anvil, and the width of the overlapping portion of the grooves is approximately double this amount, so that the wing may travel from one groove to the next, following a zigzag path relative to the hammer. The grooves 55, wings 44a and 44b, and the threads on the lead screw bushing 30, are all right-hand in the illustrative embodiment butmay be all left-hand if desired. Each of the two grooves 55 starts at an impact shoulder 56L near the center of one of the halves of the hammer and extends through the other half and up to an impact shoulder 56R. Shoulders 56R drive and impact the edges of the anvil wings when the tool is being driven in a right-hand direction, while impact shoulders 56L are arranged for driving in the left-hand direction.

For the purpose of clarifying the disclosure,

and an axial component of force, the direction of the applied force being as indicated by the arrows on the right side of Figs. 13, 14 and 15. Since the anvil and lead screw 30 are held against axial movement and the rotation of the anvil is resisted by the external load, the resultant of the forces acting ,on the hammer tends to move the sides of the anvil wing it. It will be seen that the succeeding impact shoulder 53R, Fig. 15,

is moving directly toward engagement with the upper edge of the wing a. It will be noticed also that the front edge 55F of the hammer groove, acting on'the front face of the anvil wing 44a, cams the hammer forwardly in opposition to the rearward thrust of the lead screw 30. The direction of movement of the hammer is indicated by the arrow on the surface of the hammer, while the direction of the force applied by the lead screw is shown by the arrow at the right of the figure.

Successive disengagement and reengagement of the hammer with respect to the anvil is caused by the rearward pull of the lead screw and the forwardly directed camming action of the anvil respectively. The back and forth movement of the hammer is necessarily accompanied by deceleration and acceleration respectively of the hammer with relation to the drive shaft 20 owing to the inclination of the threads on the lead screw 30. In other words, the hammer rotates slower than the drive shaft while it is disen gaging the anvil and rotates faster than the shaft 20 during the time the hammer is moving into re-engaging position relative to the anvil, the mean speed of the hammer, of course, being the same as that of the shaft but greater than the speed of the anvil. Thus, acceleration and deceleration of the hammer relative to the drive shaft are effected without necessity of a resilient connection.

The impact shoulders 56R are always in complete engagement with the edges of the anvil wings a and 4412 at the instant of impact irrespective of the angular velocity of the hammer prior to impact. It is therefore possible to rotate the drive shaft 20 at any desired speed without liability of the impact shoulders engaging when they are not in complete mesh. The force of the impact varies with the square of the difference of angular velocity between the hammer 43 and the anvil 44 at the time of impact.

The hammer delivers torque to the anvil continuously,.even during the interval between two successive impacts. This is an important feature of the invention because it obviates backlash in the connections between the toolhead and the nut or bolt which it drives. In prior, devices of this character there was lost motion between the hammer and anvil between impacts and any looseness in the connections at the wrench socket permitted the anvil to slip back during the 'lost motion interval and resulted in dissipation of a portion of the impact force before the slack was taken up. In the device of the present invention, however, the tool head is securely held in operative driving relation of the driven nut or bolts so that, at the instant of impact the momentum of the hammer is checked in a minimum time and consequently with a maximum instantaneous striking force. Referring to Figs. 14 and 15, it will be seen that the front edge 55! of the hammer groove imparts torque to the adjacent face backhead has a front wall 62 of the anvil wing a from the time that the impact shoulder 56R is disengaged from the anvil up to the time that the shoulder re-engages the anvil with an impact. It will be understood that the axial movement of the hammer is resisted by its own inertia during this interval.

While the illustrative embodiment of the hammer and anvil assembly comprises a grooved hammer surrounding the anvil, it will be understood that the arrangement may be' reversed so that the anvil is grooved and surrounds the hammer. The form shown in the drawings is preferred because it minimizes the moment of inertia or flywheel effect of the stricken element.

If the shaft 29 be driven in a counterclockwise direction, the operation is similar to that previously described. In this instance, however, the

lead screw 39 tends to push the hammer 43 forwardly instead of rearwardly as in the case-of clockwise rotation.- Moreover, the shoulders 56L drive the anvil and deliver impacts thereto and the back edges B of the hammer groove serve to impart torque to the anvil and to move the hammer rearwardly, during the interval between impacts Figs. 18 to 26 inclusive illustrate a preferred arrangement of a prime mover and power transmission for driving the shaft 29. The mechanism for this purpose is supported in a casing 69 secured by bolts 21 to the bearing support housing 24. A backhead BI is secured to casing 69 by any suitable means, such as bolts (not shown). The into which is pressed a bushing 63. The bushing supports a roller bearing 64 for the rotor shaft 65 of a rotary air motor 66. The front end of the shaft 65 is supported in a roller bearing 61. The latter is supported by means of an integral wall 68 on the casing 69, a bushing 69 and a cap 19 bolted to wall 68. I

The motor 66 comprises a cylinder II, end plates 12, a rotor I3 integral with the shaft 66; and blades M carried by the rotor. For a more complete description of a suitable type of rotary air motor, reference is made to Amtsberg application, Serial No. 35,920; filed August 13, 1935, now Patent 2,977,733, April 20, 1937.

Compressed air, or other suitable pressure fluid, is supplied to the motor 66 through an admission valve 15 in a handle I6 on the casing 69, and a passageway I8 in said casing which leads into the backhead 6|. The live air then passes through a pressure regulator or throttling governor (to be described presently), thence through passage 19 leading to the inlet port al in a valve housing 82. A reversing valve 83 is mounted for turning movement in said housing and is provided with a pair of arcuate recesses 84. Housing 92 has an exhaust port 85 and a pair of ports 86 communicating with the respective sides of the rotary motor 66. In the position shown in Fig. 24, the left side of the motor is connected to live pressure fluid and the right side to exhaust, so that the motor will run clockwise looking rearwardly. It willbe understood that by turning the reversing valve 84 through the motor may be conditioned for counterclockwise rotation. A manipulative handle 88 and a spring-pressed detent 89 I cess 96 receiving live pressure fluid through a port 91 in the sleeve 99.

The rotatable unit 93 consists of a collar having a frusto-conical head 98 and a cylindrical portion 99, and is split into segments by means of slots I99 extending from the front end of the collar almost to the rear end. The rear end of collar 93 has a running flt with the housing 92 and is secured by a pin ml to the rotating shaft 9|. Housing 92 has a frusto-conical portion I92 juxtaposed to the head 98 on collar 93 and forming therebetween an annular passageway I93 of variable thickness. The housing has a series of ports I94 supplying live pressure fluid from the recess 96 to the passageway I93.

The operation of the governor is as follows: Compressed air or other pressure fluid is admitted through the port 91 in the sleeve 99 to the annular recess 96 in housing 92. The fluid passes through holes I94 to the annular passageway I93. From this space the fluid flows out through passage I9 and through the reversing valve and motor, as previously described, to drive the shaft 69 and its extension 9I. Collar 93 rotates with shaft 95 and the sections of the head-98 are therefore thrown outwardly by centrifugal force toward the frusto-conical surface I92 on the stationary housing 92. Outward movement of the segments of the head 98 (resisted by the stiffness of the split cylindrical elements) restricts the annular passageway I 93 between the fixed and movable parts of the governor and thereby reduces the supply of pressure fluid to the rotary motor66. Since the speed of the rotary motor depends on the rate at which pressure fluid is supplied to it through passage I9, it will be seen that any variations from the normal speed of the motor will be resisted by the action of the governor.

To adjust the speed, the operator turns the knob 95, thus moving the housing 92 in an axial direction to increase or reduce the normal size of the annular passageway I93 between the head 98 and the surrounding portion of housing 92.

The power transmission between the rotary shaft 65 and the shaft 29 in the front housing 26.

provides a speed reduction, resiliency and lost motion. Shaft 20 has an integral flange I95 in the casing 69, said flange being provided with an annular gear I96. A crank shaft I91 extending parallel to the aligned shafts 65 and 29',- has an integral lateral extension I98 rotatably mounted on shaft 29 by means of a ball bearing I99. A crank arm H9, rigidly secured to and projecting laterally from the rear end of the crank shaft I91, rotates on roller bearings III mounted on cap 19 which surrounds the shaft 65.

A countershaft or cluster gear H2 is mounted to rotate on crank shaft I91. The rear end of the countershaft has gear teeth H3 meshing with a spur gear II4 on the front end of the motor shaft 65. The front end of the countershaft has gear teeth I I5 meshing with the annular gear Gears 4, II3,- H5, and I96 provide an epicyclic gearing speed reduction between rotor shaft 65 and shaft 29. It will be understood that the resistance of shaft 29 to rotation tends to cause the crank shaft I91 to revolve about the axis of the shaft 29 and 66.

Resilient means resist planetary movement of the crank shaft l 01. Such means comprise a connecting rod ll! rotatably mounted at its upper end on the countershaft I I2 and rotatably mounted at its lower end on a wrist pin I I8. The wrist pin is supported by a piston H9 mounted to reciprocate in a cylinder I20. The cylinder is in the form of a sleeve removably supported in the casing 60 and having at its lower end one or more holes |2l communicating with the live air duct 18.

As long as the admission valve 15 is held open, live air, under constant pressure, admitted to the cylinder I20 through duct 18 and openings I2l, tends to hold the piston H9 and the crank shaft I01 in their uppermost position, shown in Fig. 21. Assuming the rotor shaft to be rotating counterclockwise (looking forwardly), the gears H3 and H5 rotate clockwise about the crank shaft I01 and at the same time tend to revolve counterclockwise about the axis of shafts 20, 65, such tendency depending upon the amount of torque applied. The forces tending to cause orbital translatory movement of the countershaft H2 occur at both ends of the countershaft and are applied in the same tangential direction by virtue of the arrangement of the gearing comprising an external and an internal speed reduction. Consequently, the tangential force resultant falls at a point between the gears H3, i [5 and may be successfully opposed by the connecting rod thrust without engineering a tangential (rocking) couple affecting the countershaft.

As the resistance of shaft 20 to rotation increases, the crank shaft HI'! moves toward the position shown in full lines in Fig. 22 and the piston is forced downwardly against the air pressure in the cylinder I20. The angular displacement of the crank shaft is a function of the torque applied by the motor shaft 65, which in turn depends on the resistance of the driven shaft 20. When the load on the driven shaft is reduced (due to the disengagement of hammer 43 from anvil Ml), the crank shaft moves back to the Fig. 21 position and, in doing so, increases the speed of the driven shaft 20 over the speed imparted to it by the rotor shaft 65; and vice versa.

If excessive resistance is encountered by the driven shaft 20, the crank shaft I01 continues moving counterclockwise as the motor shaft 65 slows down, until the crank shaft has been displaced 180 away from its Fig. 21 position. At

this point the resilient coupling reverses itself and, as it continues to move counterclockwise, the piston H9 augments the motor shaft in turning the crank shaft and consequently the angular velocity of the motor shaft is increased. The latter may therefore rotate continuously without turning the driven shaft 20, the lost motion being due to the fact that the rotary motor 66 drives the piston H9 during one half of a cycle and is driven by the piston during the succeeding half of the cycle. The broken lines in Fig. 22 indicate the position of the piston and crank shaft during a part of the cycle in which the piston drives livered by the motor shaft 65. The form of the curve is substantially straight-line up to about which indicates that the eflection 9 the crank shaft varies substantially in direct proportion to the torque transmitted through the gearing within wide limits and the ratio between torque and angular deflection diminishes thereafter. The maximum transmissible torque is realized with a crank shaft deflection of or more, depending upon the ratio of crank radius to connecting rod length, and thereafter a diminishing torque serves further to deflect the crank shaft. Beyond displaceinent of the crank shaft, the torque delivered by the motor to the flexible coupling is negative in effect and the piston reverses its movement and drives the motor until the parts are restored to the Fig. 21 position.

The elastic characteristics of the coupling may be varied to suit different requirements by. altering the force acting upon the piston, as by modifying the piston area or changing the air pressure in the cylinder. An advantage lies in supplying the cylinder and the air motor from a common source of pressure fiuid, in that changes in the latter will affect both equally, thus providing a softer coupling when motor capacity is reduced, and vice versa. A mechanical compression spring may be substituted for the action of live air in the cylinder. The coupling characteristic then will be modified by the character of the spring. a

The coupling transmits positive or negative torque with identical elastic action, the crank shaft deflection-corresponding in sign. This is an advantage as no duplication of parts is entailed for reversing.

Since the preferred form of driving means for the hammer delivers a substantially constant torque thereto throughout the cycle of operation, it follows that for any particular torque selected the hammer will be driven at a definite mean R. P. M. Furthermore, since all the) principal forces involved in driving the hammer are proportional to the accelerations, it can be demonstrated that the R. P. M. of the hammer will vary as the square root of the torque applied by the rotary air motor. The impact forces vary as the square of the impact velocities, which in turn vary directly as the product of'the acceleration and the time, and it follows, therefore, that the magnitude of the blows delivered by the hammer will vary diquently, the speed regulating governor may be dispensed with, if desired, and the force of blows regulated by selecting a corresponding air pressure for driving the motor. The number of blows per minute will then depend on the fixed characteristics of the wrench, such as hammer inertia,

and will be different for varying air pressures.

What is claimed is:

1. An impact clutch comprising a rotatable hammer and a rotatable anvil, means for driving said hammer, successively disengageable and reengageable impact shoulderson said hammer and anvil respectively, automatic means for causing disengagement of said impact shoulders, and means for automatically effecting re-engagement of said shoulders, characterized in that the re-engaging means comprises complementary cam surfaces on said hammer and anvil.

2. An impact clutch comprising a rotatable hammer, a rotatable anvil, complementary impact shoulders on said hammer and anvil separable by relative axial movement between the hammer and anvil, driving means for the hammer, automatic means for disengaging the impact shoulders to permit rotation of the hammer ahead of the anvil, and automatic means for effecting re-engagement of the impact shoulders, said re-engaging means comprising complementary engageable cam surfaces carried by the hammer and anvil respectively.

3. An impact clutch comprising a hammer eleelements adapted for the transmission of torque,

said shoulders being disengageable and re-engageable upon relative axial movement of said elements, and complementary interengaging means carried by said elements for causing the driving element to follow a predetermined path relative to the driven element during the time said shoulders are out of engagement.

5. A clutch comprising a driving element, a driven element, complementary shoulders on said elements adapted for the transmission of torque, said shoulders being disengageable and re-engageable upon relative axial movement of said elements, one of said elements having guiding means carried thereby and received within a groove formed in the other element, said groove being of helical shape from one shoulder to the succeeding shoulderio compel alignment of the driving and driven shoulders upon reengagement.

6. An automatic clutch comprising a driving element and a driven element, one of said elements comprising a stem and a flange, the other element having disk-like portions engaging opposite faces of said flange in alternation and forming a zigzag groove receiving said flange, said flange and grooved portion having complementary disengageable and reengageable shoulders for the transmission of torque.

7. An impact clutch comprising a rotatable shaft, a rotatable impact hammer element, a hellcal driving connection between the shaft and hammer element permitting relative rotary and axial movements therebetween, an anvil element, one of said elements comprising a plate having clutch teeth projecting in opposite directions and the other element comprising a pair of rigidly connected and axially spaced parts, one on each side of the first-mentioned plate and having clutch teeth alternatively engageable with the first-mentioned clutch teeth.

10. An impact clutch comprising a rotatable shaft, a rotatable impact hammer element, a

7 driving connection between the shaft and hammer element, an anvil element, one of said elements having axially spaced sets of clutch teeth, complementary sets of teeth on the other element, a driving connection having torque responsive means tending to move the hammer element axially in one direction or the other depending upon the direction of rotation, and cam means between said clutch elements for moving the hammer element axially relative to the anvil element oppositely and subsequently to the axial movement caused by said torque responsive means.

11. An impact tool comprising a drive shaft, a rotatable impact hammer having a screwthreaded connection with said shaft, a rotatable anvil held against axial movement relative to the shaft, said hammer and anvil having complementary impact shoulders separable by axial movement of the hammer, said screw-threaded connection tending to separate said shoulders, and means for automaticallycausing re-engagement oi?- the shoulders following separation, said lastmentioned means comprising complementary cam surfaces carried by the hammer and anvil respectively.

12. An impact tool comprising a rotatable impact hammer, means for driving said hammer, a rotatable anvil, said hammer and anvil having successively engageable and disengageable impact shoulders separable by relative axial movement between the hammer and anvil, means including said shoulders for causing the hammer to deliver rota y mpacts to the anvil, and means for driving said anvil while the shoulders are out of engagement.

MILLER R. HUTCHISON, JR. 

